Rotary mechanism control system with bilateral braking

ABSTRACT

The invention relates to a hydraulic controller (1), in particular for the control of the rotating mechanism of an excavator. A hydraulic motor (5) is driven by means of a hydraulic pump (2) in a hydraulic drive circuit, the hydraulic pump (2) and the hydraulic motor (5) being connected by a first and a second working line (3, 4). The hydraulic controller (1) includes an adjustment arrangement (26) for adjusting a setting piston (28a), acting on the displacement volume of the hydraulic pump (2), arranged between two setting pressure chambers (26, 27), in dependence upon the pressure difference between two control lines (20, 21). In accordance with the invention, a respective separate brake valve (24; 25) is provided in each connection between each of the two setting pressure chambers (26; 27) with the pressure fluid tank (10).

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a hydraulic controller in particularfor the control of the rotating mechanism of an excavator.

2. Discussion of the Prior Art

A hydraulic controller such as for the control of the rotating mechanismof an excavator is known from DE 44 05 472 A1. This publicationdiscloses a hydraulic controller for the control of a rotating mechanismhaving a brake valve. The brake valve serves for the sensitive brakingof the rotating mechanism by means of control of the braking moment. Thebrake valve connects a pre-control arrangement, controlling the settingarrangement, with the pressure fluid tank. Thereby, the return flow ofthe pressure fluid, out of the setting pressure chamber acted on uponacceleration of the rotating mechanism, to the pressure fluid tank, isthrottled during the braking, and thus the braking procedure is delayed.

Disadvantageous with this known hydraulic rotating mechanism controlleris, however, that solely one brake valve is provided and thus the returnflow of the pressure fluid out of the two setting pressure chambers ofthe setting arrangement is not effected independently of each other.This can influence the dependability of the hydraulic controller.Further, with this known hydraulic rotating mechanism controller, it isdisadvantageous that the brake valve responds also when the rotationmovement of the rotating mechanism encounters a resistance upon braking.Such a resistance is e.g. brought about in that the excavator is locatedon an inclined plane and the boom of the rotating mechanism moves uphillduring the braking procedure. This situation occurs relativelyfrequently on building sites which naturally have uneven ground.Further, a corresponding resistance arises when the boom of the rotatingmechanism swings into a mound of earth or the like.

SUMMARY OF THE INVENTION

It is thus the object of the present invention to so further develop theknown rotating mechanism controller that dependability is increased.

The solution in accordance with the invention is based upon the insightthat the dependability of the hydraulic controller can be increased inthat its own, separate brake valve is associated with each of the twosetting pressure chambers of the adjustment arrangement, whereby each ofthe two brake valves is controlled by means of the force differencebetween a setting force exercised by the control pressure and a returnforce exercised by a return member.

A further solution pursuant of the invention is based on the insightthat in the working lines connecting the hydraulic pump with thehydraulic motor a pressure reversal occurs when the rotating mechanismcan rotate further without resistance during the braking procedure.When, however, during the braking procedure the rotating mechanism isexposed to a resistance, e.g. due to the down-slope force or an impact,this pressure-side reversal does not occur, i.e. the working line actedupon with high pressure during they acceleration phase is also actedupon with high pressure during the braking procedure. The inventionexploits this insight in that for each of the setting pressure chambersa respective separate brake valve is provided which in each case isconnected with one of the working lines. The brake valves are therebycontrolled in dependence upon the pressure difference between theworking pressure in the associated working line and the controlpressure. Thereby during the braking procedure with substantiallypressureless control lines, only that brake valve responds theassociated working line of which is acted upon with high pressure.

The brake valves may be constituted as switch-over valves having athrottled and a non-throttled switching position. During theacceleration phase, in which the control lines are acted upon withcontrol pressure, the brake valves are in their non-throttled switchingposition. During the braking phase, in which the control lines aresubstantially pressureless, the brake valves are switched into theirthrottled switching position, in order to delay the braking procedure.

Corresponding to an embodiment, the brake valves can have each a controlpressure chamber which is connected with the control lines. Forselection of that control line which carries the higher pressure achange-over valve may be provided.

Corresponding to another aspect, each brake valve is connected with thatworking line which, in the swinging out of the hydraulic pump prior tothe braking procedure, forms the low pressure return line of the drivecircuit.

The brake valves may be formed as switch-over valves having a throttledand a non-throttled switching position. During the acceleration phase,in which the control lines are acted upon with control pressure, thebrake valves are located in their non-throttled switching position.During the braking phase, in which the control lines are substantiallypressureless, that brake valve whose associated working line is actedupon with high pressure is switched into its throttled switchingposition.

In the case of resistance-free braking of the rotating mechanism apressure-side reversal occurs in the working line, i.e. that workingline which formed the low pressure working line during the accelerationphase, provides the high pressure working line in the braking phase.Thus, that brake valve responds whose associated setting pressurechamber was acted upon with setting pressure during the accelerationphase. In contrast, if the rotating mechanism is exposed to resistanceduring the braking procedure, this pressure-side reversal does not tookplace. The brake valve associated with that setting pressure chamberwhich was acted upon with setting pressure during the acceleration phasedoes not respond in this case, so that the corresponding settingpressure chamber can be rapidly relieved via the brake valve in thenon-throttled switching position. Thereby, an uncontrolled furtherslewing of the rotating mechanism is prevented.

Furthermore, the brake valves may have each two control pressurechambers, whereby one of the control pressure chambers is connected withthe associated working line and the other control pressure chamber withthe control lines. For the selection of that control line which carriesthe higher pressure there may be provided a change-over valve.

According to the invention, the brake valves can be particularlyadvantageously arranged directly at the control lines, wherebycorresponding to claim 12 there may be provided between the brake valvesand the associated setting pressure chambers an after-suctionarrangement, in order to ensure a prompt after-flow of pressure fluid onthe suction side upon return of the setting piston.

For limiting the control pressure in the control lines to a maximumpressure, a pressure cut-off valve may be provided.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will be described in more detail below with reference to apreferred exemplary embodiment and with reference to the drawings. Inthe drawings:

FIG. 1 shows a first exemplary embodiment of the hydraulic controller inaccordance with the invention, in the neutral position,

FIG. 2 shows the exemplary embodiment according to FIG. 1, during theacceleration phase,

FIG. 3 shows the exemplary embodiment according to FIG. 1 during thedelay phase, when the rotating mechanism experiences no rotationresistance,

FIG. 4 shows the exemplary embodiment according to FIG. 1 during thedelay phase, when the rotating mechanism experiences a rotationresistance,

FIG. 5 shows a second exemplary embodiment of the hydraulic controllerin accordance with the invention, in the neutral position, and

FIG. 6 shows a third exemplary embodiment of the hydraulic controller inaccordance with the invention, in the neutral position.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

FIG. 1 shows a first exemplary embodiment of the controller inaccordance with the invention. In the exemplary embodiments, thecontroller 1 is configured for the control of the rotating mechanism ofan excavator.

The adjustable hydraulic pump 2 is connected via a drive shaft 50 with adrive motor (not shown), e.g. a Diesel motor. By way of a drive circuitformed by the working lines 3 and 4, the hydraulic motor 5 is inconnection with the hydraulic pump 2. The hydraulic motor 5 drives, viaa drive shaft 6, the rotating mechanism (not shown) of the excavator.

The pressure fluid in the drive circuit is after-fed via a feedarrangement 7, which includes a feed pump 8 likewise connected with thedrive motor. The feed pump 8 continuously sucks pressure fluid, via afeed filter 9, out of a pressure fluid tank 10, and feeds this into thefeed line 11. The feed line is connected with the working lines 3 and 4via check valves 12 and 13, and feeds the pressure fluid in each caseinto the working line 3 or 4 carrying the lower pressure. The feedpressure is regulated by means of pressure regulation valves 14 and 15.The over-pressure valve 16 serves for limiting the pressure in the feedline

The control of the displacement volume of the hydraulic pump 2 iseffected via the hand controller 17, which is in connection with thepressure fluid tank 10 and, via the control pressure filter 18, with acontrol pressure in-feed 19. The hand controller 17 applies a controlpressure to one of the two control lines 20 or 21 in dependence upon itsdeflection. In the neutral position illustrated in FIG. 1, both controllines 20 and 21 are vented via the hand controller 17 to the pressurefluid tank 10.

The control lines 20, 21 are connected each with a respective settingpressure chamber 26 and 27 of the adjustment arrangement 28, viathrottle points 22 and 23 and brake valves 24 and 25 to be described inmore detail.

Between the setting pressure chambers 26 and 27, there is arranged asetting piston 28a, which adjusts the displacement volume of thehydraulic pump 2 via a piston rod 29. The setting piston 28a is centeredin its neutral position shown in FIG. 1, via centering springs 30 and31.

Between each brake valve 24 and 25 and the setting pressure chamber 26and 27 associated therewith, an after-suction arrangement 32 opens intothe control lines 20 and 21. In the illustrated exemplary embodiment,the after-suction arrangement 32 consists of two check valves 33 and 34and serves for the after-suction of pressure fluid out of the pressurefluid tank 10 during the return of the setting piston 28a into itsneutral position.

The brake valves 24 and 25 have each two control pressure chambers 35and 36 or 37 and 38. The control pressure chambers 36 and 37 areconnected with the two control lines 20 and 21 via a change-over valve39. The control pressure chambers 35 and 38 lying opposite to thecontrol pressure chambers 36 and 37 are connected each with one of thetwo working lines 3 or 4 via working line connection lines 40 and 41. Inthe neutral position shown in FIG. 1, both control lines 20 and 21 arevented to the pressure fluid tank 10 via the hand controller 17, so thatthe control pressure chambers 36 and 37 are pressureless. Since thesetting piston 28a is located in its neutral position and thus thehydraulic pump works with zero displacement volume, the working lines 3and 4 are likewise pressureless so that no pressure difference arisesbetween the control pressure chambers 35 and 37 on the one hand and 36and 38 on the other hand. Thus, the brake valves 24 and 25 are heldtheir non-throttled switching positions 44 and 45 by means of theadjustable pressure springs 42 and 43.

For limiting the control pressure in the control lines 20 and 21, afterthrottle points 22 and 23 there is provided in the exemplary embodimenta pressure cut-off valve 47 which can be set by means of a setter 46,which limits the pressure-carrying control line 20 or 21, upon apredetermined maximum pressure being exceeded, to the pressure fluidtank 10. A further over-pressure valve 48 is controlled by the workinglines 3 and 4 via a change-over valve 49.

The functioning of the brake valves 24 and 25 of the first exemplaryembodiment, in accordance with the invention, will be described in moredetail below with reference to the operating conditions of the hydrauliccontroller illustrated in FIGS. 2 to 4.

FIG. 2 shows the hydraulic controller in the acceleration phase. Bymeans of deflection of the control stick 60 of hand controller 70, thecontrol line 20 is acted upon with control pressure via the controlpressure filter 18 from the control pressure in-feed 19, whilst theother control line 21 is vented to the pressure fluid tank 10. Thereby,the setting pressure chamber 26 is acted upon with control pressure viathe brake valve 24, so that the setting piston 28a displaces in thedirection indicated by the arrow 61. The hydraulic pump 2 iscorrespondingly swung out, and a corresponding high pressure built up inthe working line 4 in order to drive the hydraulic motor 5 in thedesired direction of rotation. In this manner, the rotating mechanism ofthe excavator coupled to the hydraulic motor 5 is accelerated. The brakevalves 24 and 25 are, thereby, located in the non-throttled switchingpositions 44 and 45, since during the acceleration phase shown in FIG. 2there is present in the control line 20 and thus also in the controlpressure chambers 36 and 37 a corresponding control pressure which urgesthe brake valves 24 and 25 into their non-throttled switching position.

After attainment of the desired rotational speed, the control stick 60can be released by the operator, so that this swings back into theneutral position illustrated in FIG. 3. Thereupon, the control line 20and also the control line 21 are vented to the pressure fluid tank 10and the control pressure in the control line 20 is reduced.Correspondingly, the control pressure chambers 36 and 37 of the brakevalves 24 and 25 are no longer acted upon with control pressure.

Because of the pressure fluid filled in the setting pressure chamber 26during the acceleration phase, the hydraulic pump 2 is however initiallystill in its swung out position. Insofar as the rotating mechanismcoupled to the hydraulic motor 5 can rotate freely in this delay phaseillustrated in FIG. 3, without being exposed to any resistance, therebuilds up a pressure in the working line 3 whilst the pressure in theworking line 4 falls below the pressure prevailing in the working line3. There thus occurs a pressure-side reversal, whereby the working line3, working as low pressure working line during the acceleration phase,now becomes the high pressure working line and the working line 4,serving as high pressure working line during the acceleration phase, nowbecomes the low pressure working line. The present invention exploitsthis effect.

The control pressure chamber 35 of the brake valve 24 connected via theworking line connection line 40 with the high pressure working line 3now brings about a switch-over of the brake valve 24 into the throttledswitching position 70. In the delay phase illustrated in FIG. 3, thesetting piston 28a is urged back into its neutral position illustratedin FIG. 1 by means of the centering springs 26 and 27, as is indicatedby means of the arrow 72. The return flow of the pressure fluid out ofthe setting pressure chamber 26 via the control line 20 towards thepressure fluid tank 10 is however throttled by the brake valve 24located in the control line 20, since this brake valve 24 is in itsthrottled switching position. The return of the setting piston 28 thusoccurs in this operational condition relatively slowly, which manifestsitself in a sensitive, delayed braking of the rotating mechanism.

The under-pressure which arises in the setting pressure chamber 27through the return movement of the setting piston 28 brings about anafter-suction of pressure fluid out of the pressure fluid tank 10 viathe after-suction arrangement 32. Thereby, the check valve 33 opens.

In FIG. 4, the operational condition in the delay phase is illustratedfor the case that the rotating mechanism is exposed to a resistanceduring the delay phase and the rotating mechanism cannot freely rotatecorresponding its moment of inertia. Such a resistance is brought aboutin particular in that the excavator (in which the rotating mechanism islocated) is standing on an inclined plane, as arises frequently onbuilding sites. When the rotating mechanism is subjected to braking in arange of angle of rotation in which the boom located on the rotatingmechanism moves up-slope, due to the down-slope forces a relativelyrapid braking of the rotating mechanism occurs. The pressure-sidereversal described with reference to FIG. 3 thereby does not occur.Rather, the hydraulic motor 5 is further driven by the hydraulic pump 2which is still swung out. In this operational condition there is thusbuilt up in the working line 4 a high pressure, whilst the working line3 works as low pressure return flow line.

In the operational condition illustrated in FIG. 4, not the brake valve24 but the brake valve 25 in this case connected with the working line 4acted upon by high pressure is displaced into its throttled switchingposition 71. It is significant for the invention that the switch-overvalve 24, in contrast to the operating condition explained withreference to FIG. 3, remains in its non-throttled switching position 44.The pressure fluid filled into the setting pressure chamber 26 duringthe acceleration phase can, during the return of the setting piston 28ainto its neutral position by means of the centering springs 30 and 31,thus escape relatively rapidly via the non-throttled brake valve 24 andthe control line 20 towards the pressure fluid tank 10.

The switching position of the other brake valve 25 is in thisoperational condition of no significance, since the pressure fluidafter-flowing into the setting pressure chamber 27 is drawn not via thebrake valve 25 but via the after-suction arrangement 32, i.e. via theopened check valve 33, out of the pressure fluid tank 10. The settingpiston 28a is thus, in contrast to the operational condition illustratedin FIG. 3, relatively rapidly returned into its neutral positionillustrated in FIG. 1. This determines a relatively rapid swinging backof the hydraulic pump 2 to zero displacement volume, so that thehydraulic motor 5 is not further driven by this pump. By these means, anuncontrolled further slewing of the rotating mechanism is effectivelyprevented in this operational condition.

This is in particular of substance also if the boom of the rotatingmechanism meets a rigid resistance, in which the boom impacts on a moundof earth or the like.

When the hydraulic pump 2 is swung out in the reverse deliverydirection, in that during the acceleration phase the control line 21 isacted upon with control pressure by means of the hand controller 17, thehydraulic controller in accordance with the invention works inaccordance with the same principle. However, in this case, the returnflow of the pressure fluid out of the setting pressure chamber 27 filledduring the acceleration phase, towards the pressure fluid tank 10, iscontrolled via the brake valve 25 in the delay phase as described above,whilst the switching position of the brake valve 24 is then of nosignificance.

FIG. 5 shows a second exemplary embodiment of the controller inaccordance with the invention, in its neutral position. Also in thisexemplary embodiment, the controller 1 is configured for the control ofthe rotating mechanism of an excavator. Elements already described areprovided with corresponding reference signs so that in this regard arepeated description is not necessary.

In the exemplary embodiment illustrated in FIG. 5, the brake valves 24and 25 have each a control pressure chamber 36 or 37. The controlpressure chambers 36 and 37 are connected via a change-over valve 39with the two control lines 20 and 21. Lying opposite the controlpressure chambers 36 and 37, respective return members are provided inthe form of return springs 80 and 81. Each of the two brake valves 24and 25 is thus controlled through the force difference between a settingforce exercised by the control line 20 or 21 which is acted upon by thehighest control pressure and a return force exercised by the respectivereturn spring 80 or 81. In the neutral position illustrated in FIG. 5,the two control lines 20 and 21 are vented to the pressure fluid tank 10via the hand controller 17, so that the control pressure chambers 36 and37 are pressureless. The braking valves 24 and 25 are thus switched intotheir respective throttled switching position 70 and 71 by means of thereturn springs 80 and 81.

Below, the functioning of the brake valves 24 and 25 in accordance withthe invention, according to the exemplary embodiment illustrated in FIG.5, will be described in more detail.

For accelerating the rotating mechanism of the excavator, the hydraulicpump 2 is swung out in one of its delivery directions corresponding tothe intended direction of rotation. For this purpose, via the controlstick 60 of the hand controller 17, the control line 20 or the controlline 21 is acted upon with control pressure, via the control pressurefilter 18, from the control pressure in-feed 19, whilst in each case theother control line is vented to the pressure fluid tank 10. Throughthis, the setting pressure chamber 26 or the setting pressure chamber 27is acted upon with control pressure via the brake valve 24 or the brakevalve 25, so that the setting piston 28a is correspondingly displaced.The hydraulic pump 2 is correspondingly swung out, and a correspondinghigh pressure built up in one of the working lines 3 or 4 in order todrive the hydraulic motor (not shown in FIG. 5) in the desired directionof rotation and to accelerate the rotating mechanism of the excavator.The brake valves 24 and 25 are in the non-throttled switching position44 and 45, since during the acceleration phase there is present in oneof the two control lines 20 or 21 and thus also in the control pressurechambers 36 and 37 a corresponding control pressure, which urges thebrake valves 24 and 25 into their non-throttled switching positions 44and 45.

After attainment of the desired speed of rotation, the control stick 60can be released by the operator, so that this swings back into itsneutral position. In response, the control line 20 and also the controlline 21 is vented to the pressure fluid tank 10 and the control pressurein the control lines 20, 21 falls. Correspondingly, the control pressurechambers 36 and 37 of the brake valves 24 and 25 are no longer actedupon with control pressure.

Due to the pressure fluid filled, during the acceleration phase, in oneof the control pressure chambers 26 or 27, the hydraulic pump 2 ishowever initially further in its swung out position. The pressure fluidescapes from the setting pressure chamber 26 or 27, acted upon withpressure fluid during the acceleration phase, via the brake valve 24 or25 associated with the setting pressure chamber and the and controller17, to the pressure fluid tank 10. Thereby, he two brake valves 24 and25 are now located in their throttled switching position 70 and 71,since these are acted upon by the associated return springs 80 and 81and the control pressure chambers 36 and 37 are substantiallypressureless. The return flow of the pressure fluid out of therespective setting pressure chamber 26 or 27 is thus throttled by therespective associated brake valve 24 or 25. The return of the settingpiston 28 thus occurs relatively slowly, which is manifested in asensitive, delayed braking of the rotating mechanism.

The under-pressure arising, due to the return of the is setting piston28, in the setting pressure chamber not acted upon with setting pressureduring the acceleration phase brings about an after-suction of pressurefluid out of the pressure fluid tank 10 via the after-suctionarrangement 32. Thereby, the respective non-return valve 33 or 34 opens.

Through the association of a respective separate brake valve 24 or 25with each setting pressure chamber 26 and 27 of the adjustmentarrangement, the operational safety of the rotating mechanism controlleris significantly improved, without substantially increasing the outlay.Through the arrangement of the brake valves 24 and 25 directly at thecontrol lines 20 and 21, a particularly rapid response of the brakevalves 24 and 25 is attained.

Further, the invention can also be employed in combination with apre-controller such as is known in principle from DE 44 05 472 A1.

FIG. 6 shows a hydraulic control in accordance with the presentinvention having another, expedient exemplary embodiment for apre-controller. The third exemplary embodiment illustrated in FIG. 6 issimilarly configured to the second exemplary embodiment illustrated inFIG. 5. Elements already described are thereby provided withcorresponding reference signs, so that the following description relatessolely to the differences and special features.

The feed arrangement 7 serves, in the exemplary embodiment illustratedin FIG. 6, not only for the after-feed of the working circuit 2 to 4 butalso for the delivery of pressure fluid to the adjustment arrangement28. Respective pressure regulation valves 90, 91 are associated with thebrake valves 24 and 25, which pressure regulation valves are eacharranged upstream of the associated brake valve 24, 25. The pressureregulation valves 90 and 91 are connected on the one hand with the feedline 11 of the feed arrangement 7 and on the other hand with thepressure fluid tank 10. Each pressure regulation valve 90, 91 isconnected via a connection line 92, 93 with the associated brake valve24, 25. The control of the pressure regulation valve 90 or 91 iseffected proportionally to the pressure difference between the settingpressure prevailing in the respective connection line 92 or 93 and thecontrol pressure in the associated control line 20 or 21. For thispurpose, in each case one of the control inputs of the pressureregulation valve 90 or 91 is connected via an associated by-pass line 94or 95 with the connection line 92 or 93. In each case another controlinput of the pressure regulation valve 90 or 91 is connected with theassociated control line 20 or 21. The setting pressure prevailing in theconnection lines 92 and 93 is thus in substance proportional to thecontrol pressure prevailing in the associated control line 20 or 21.However, by means of the pressure springs 96 and 97 it is attained thatthe setting pressure is slightly, e.g. 1 to 2 bar, above the controlpressure prevailing in the associated control line 20 or 21.

During the acceleration phase the setting pressure prevailing in therespective setting pressure chamber 26 or 27 is metered through therespective pressure regulation valve 90 or 91 in substanceproportionally to the control pressure prevailing in the respectivecontrol line 20 or 21. In the delay phase, the pressure fluid flows outof the setting pressure chamber 26 or 27, in the manner described withreference to FIGS. 1 to 4, via the respective brake valve 24 or 25 backto the pressure fluid tank 10 via the respective pressure limiting valve90 or 91.

The after-suction arrangement 32 is not necessary in this exemplaryembodiment since the pressure fluid supply of that setting pressurechamber the volume of which increases in the return to the neutralposition is effected via the feed arrangement 7, the feed line 11 andthe associated pressure regulation valve 90 or 91, and the associatedbrake valve 24 and 25. The advantage lies in particular in that nosuction resistance is to be overcome; rather, via the feed pump 8, anactive feed into the adjustment arrangement 28 is effected. A possiblecontamination by means of dirt particles is reliably and effectivelyavoided by means of the feed filter 9. In principle it is also possibleto provide an after-suction filter in the passive after-suctionarrangement 32. However, in comparison with the feed filter 9, this mustbe substantially larger in order to maintain the suction resistance aslow as possible. This, however, conflicts with the goal of aconstruction which is as compact as possible.

The invention is not limited to the illustrated exemplary embodiment. Inparticular, the brake valves 24 and 25 need not necessarily be arrangeddirectly in the control lines 20 and 21. They may be provided at anyposition in the return flow line between the setting pressure chambers26 and 27 and the pressure fluid tank.

What is claimed is:
 1. Hydraulic controller, for the control of therotating mechanism of an excavator, havinga hydraulic drive circuit(2-4) with a hydraulic pump (2) and a hydraulic motor, and a first and asecond working line (3,4) connecting the hydraulic pump (2) with thehydraulic motor, an adjustment arrangement (26) for adjusting a settingpiston (28a), acting on the displacement volume of the hydraulic pump(2), arranged between two setting pressure chambers (26,27), independence upon the pressure difference between two control lines(20,21) and characterized in that,a respective separate brake valve(24;25) is associated with each connection between teach of the twosetting pressure chambers (26;27) with a pressure fluid tank (10), eachsaid respective brake valve, responsive to a disappearing pressure inthe control lines (20,21), throttling a return flow of the pressurefluid out of the setting pressure chambers (26,27) into the fluidpressure tank (10), whereby each of the two brake valves (24;25) iscontrolled by means of a force difference between a setting forceexercised by the control pressure in the control line (20,21) acted uponby the greater control pressure and a return force exercised by a returnmember (80;81) provided for each brake valve (24;25).
 2. Hydrauliccontroller according to claim 1,characterized in that,each brake valve(24; 25) is formed as a switch-over valve having a first switchingposition (44; 45) with non-throttled through-flow and a second switchingposition (70;71) with throttled through-flow, whereby each brake valve(24; 25) is in the first switching position (44; 45) when the forcedifference between the setting force and the return force is greaterthan a predetermined threshold value and each brake valve (24; 25) is inthe second switching position (70; 71) when the force difference betweenthe setting force and the return force is smaller than the predeterminedthreshold value.
 3. Hydraulic controller according to claim 1 or2,characterized in that,each brake valve (24; 25) has a respectivecontrol pressure chamber (36; 37) which is connected with the is controllines (20; 21).
 4. Hydraulic controller according to claim3,characterized in that,the control pressure chambers (36; 37) of thebrake valves (24; 25) are connected with the control lines (20; 21) viaa change-over valve (39).
 5. Hydraulic controller according to claim1,characterized in that,the return members are formed as return springs(80, 81).
 6. Hydraulic controller, for the control of the rotatingmechanism of an excavator, havinga hydraulic drive circuit (2-4) with ahydraulic pump (2) and a hydraulic motor (5), and a first and secondworking line (3,4) connecting the hydraulic pump (2) with the hydraulicmotor (5), an adjustment arrangement (28) for adjusting a setting piston(28a), arranged between two setting pressure chambers (26,27), acting onthe displacement volume of the hydraulic pump (2), in dependence uponthe pressure difference between two control lines (20,21) and arespective separate brake valve (24;25) is associated with eachconnection between each of the two setting pressure chambers (26;27)with a pressure fluid tank (10), each said respective brake valve,responsive to a disappearing pressure in the control lines (20,21),throttling a return flow of the pressure fluid out of the settingpressure chambers (26,27) into the fluid pressure tank (10), whereby afirst of the two brake valves (24) is controlled by means of thepressure difference between the working pressure in the first workingline (3) and the control pressure in the control line (20,21) acted uponby the greater pressure, and the second of the two brake valves (25) iscontrolled by means of the pressure difference between the workingpressure in the second working line (4) and the control pressure in thecontrol line (20,21) acted upon by the grater pressure.
 7. Hydrauliccontroller according to claim 6,characterized in that,each brake valve(24; 25) is respectively connected with that working line (3; 4) whichupon swinging out of the hydraulic pump (2) by means of action upon thesetting pressure chamber (26; 27) associated with the brake valve (24;25) forms the low pressure return line (3; 4) of the drive circuit(3,4).
 8. Hydraulic controller according to claim 6 or 7,characterizedin that,each brake valve (24; 25) is formed as a switching valve havinga first switching position (44; 45) with non-throttled through-flow anda second switching position (70; 71) with throttled throughflow wherebythe brake valve (24; 25) is in the first switching position (44; 45)when the pressure difference between the working pressure in theassociated working line (3; 4) and the control pressure in the controlline (20, 21) acted upon with the greater pressure is smaller than apredetermined threshold value, and the brake valve (24; 25) is in thesecond switching position (70; 71) when the pressure difference betweenthe working pressure in the associated working line (3; 4) and thecontrol pressure in the control line (20, 21) acted upon with thegreater pressure is greater than the predetermined threshold value. 9.Hydraulic controller according to claim 6,characterized in that,eachbrake valve (24; 25) has two control pressure chambers (35, 36; 37, 38),whereby a first control pressure chamber (35; 38) is connected with theassociated working line (3; 4) and the second control pressure chamber(36; 37) is connected with the control lines (20, 21).
 10. Hydrauliccontroller according claim 9,characterized in that,the two controlpressure chambers (36, 37) are connected with the control lines (20, 21)via a change-over valve (39).
 11. Hydraulic controller according toclaim 1 or 6,characterized in that,the brake valves (24, 25) arearranged in the control lines (20, 21).
 12. Hydraulic controlleraccording claim 11,characterized in that,an after-suction arrangement(32), for the after-suction of pressure fluid out of the pressure fluidtank (10), is provided between the brake valves (24; 25) and theassociated setting pressure chambers (26, 27).
 13. Hydraulic controlleraccording to claim 1 or 6,characterized in that,the control lines (20,21) can be alternately acted upon with control pressure or vented to thepressure fluid tank (10) via a controller (17) connected with thepressure fluid tank (10) and a control pressure in-feed (19). 14.Hydraulic controller according to claim 1 or 6,characterized in that,apressure cut-off valve (47) is provided for limiting the controlpressure in the control lines (20, 21) to a maximum pressure.